Control device of spark-ignition engine

ABSTRACT

A control device of a spark-ignition engine is provided. The control device includes a main body of the engine, a fuel injection valve, an ignition plug, and a controller. According to an operating state of the engine, the controller switches an ignition mode between a compression-ignition mode in which compression-ignition combustion is performed by causing the mixture gas to self-ignite and combust, and a spark-ignition mode in which spark-ignition combustion is performed by igniting the mixture gas with the ignition plug to combust. The controller switches the ignition mode from the spark-ignition mode to the compression-ignition mode by performing a transition mode in which a temperature inside the cylinder is forcibly decreased and combustion is performed.

BACKGROUND

The present invention relates to a control device of a spark-ignitionengine.

Prior art approaches for improving both an exhaust emission performanceand a thermal efficiency are known. For example, JP2007-154859Adiscloses using a combustion mode of compressing to ignite mixture gasinside a cylinder. However, compression-ignition combustion occurs witha significant increase in pressure as the engine load increases, causingan increase in combustion noises. Thus, as disclosed in JP2007-154859A,even with engines which perform compression-ignition combustion, withinan operating range on a high engine load side, instead ofcompression-ignition combustion, spark-ignition combustion by operationof an ignition plug has generally been performed.

JP2009-197740A discloses an engine which performs compression-ignitioncombustion within a low engine load operating range with low enginespeed similarly to the engine disclosed in JP2007-154859A. With theengine, within the compression-ignition combustion performing range,open periods of intake and exhaust valves are adjusted to leave burnedgas at a high temperature inside a cylinder so that the in-cylindertemperature is increased to stimulate the compression-ignitioncombustion. Whereas, within a part of the compression-ignitioncombustion performing range where the engine load is relatively high andthe engine speed is relatively high, the open timing of an intake valveis advanced so that burned gas inside the cylinder is blown back to theintake port side once and then introduced into the cylinder again alongwith fresh air. In this manner, the temperature of the burned gasdecreases because of the fresh air, and thus, within the relativelyhigh-engine-speed high-engine-load range where the temperature and thepressure at the end of compression stroke become high, the significantpressure increase due to the compression-ignition combustion can besuppressed.

Meanwhile, in spark-ignition combustion, since thermal efficiency isrelatively low, the combusting gas temperature increases. Whereas, incompression-ignition combustion, as described in JP2007-154859A andJP2009-197740A, the high-temperature burned gas is introduced into thecylinder to secure the ignitability. Therefore, with engines in whichthe combustion mode is switched according to the engine operating state,a temperature environment inside the cylinder is comparatively high andthe high-temperature burned gas produced by the spark-ignitioncombustion is introduced into the cylinder immediately after thespark-ignition combustion is switched to the compression-ignitioncombustion, resulting in an excessive increase in the in-cylindertemperature. This excessive increase may cause pre-ignition such thatthe mixture gas within the cylinder is compressed to ignite in, forexample, a compression-stroke period, and a pressure increase rate(dP/dθ) inside the cylinder may become significantly high to cause loudcombustion noises.

Note that switching from spark-ignition combustion tocompression-ignition combustion is not limited to be performed accordingto the engine load decrease, but may also be performed while the engineload is stable, as well as when an engine temperature increases from acold-start state to a warmed-up state and under other circumstances.

SUMMARY

The present invention is made in view of the above situations and avoidsan increase in combustion noises when switching from spark-ignitioncombustion to compression-ignition combustion.

In the present invention, when switching from spark-ignition combustionto compression-ignition combustion, a transition mode for performingcombustion in which an in-cylinder temperature is forcibly decreasedintervenes therebetween. Thus, the temperature of the exhaust gasintroduced into the cylinder immediately after the combustion isswitched to the compression-ignition combustion is decreased to reducethe in-cylinder temperature, and therefore, pre-ignition of the mixturegas is avoided to prevent generation of combustion noises.

According to one aspect of the invention, a control device of aspark-ignition engine is provided. The control device includes a mainbody of the engine having a cylinder, a fuel injection valve forinjecting fuel to be supplied into the cylinder, an ignition plug forigniting mixture gas within the cylinder, and a controller for operatingthe engine by controlling at least the fuel injection valve and theignition plug.

According to an operating state of the engine, the controller switchesan ignition mode between a compression-ignition mode in whichcompression-ignition combustion is performed by causing the mixture gasto self-ignite and combust, and a spark-ignition mode in whichspark-ignition combustion is performed by igniting the mixture gas withthe ignition plug to combust the mixture gas. The controller switchesthe ignition mode from the spark-ignition mode to thecompression-ignition mode by performing a transition mode in which atemperature inside the cylinder is forcibly decreased and combustion isperformed.

Here, “switching the ignition mode (between the compression-ignitionmode and the spark-ignition mode) according to the operating state ofthe engine” includes a case of switching between thecompression-ignition mode and the spark-ignition mode according to achange of an engine load, a case of switching from the spark-ignitionmode to the compression-ignition mode due to a change of an enginetemperature state changing from either one of a cold-start and awarming-up state where the engine temperature is below a predeterminedvalue, to a warmed-up state, a case of switching to thecompression-ignition mode due to the engine in the spark-ignition modeof an idle state shifting an operating range to other than the idlestate, and a case of switching to the normal compression-ignition modeafter resuming from a deceleration fuel cut to the spark-ignition modewhen starting the fuel supply. Moreover, there is also a case ofswitching between the compression-ignition mode and the spark-ignitionmode without any substantial change of the engine load.

With the above configuration, when switching from the spark-ignitionmode to the compression-ignition mode, the transition mode is performed.The transition mode is a mode in which the temperature inside thecylinder is forcibly decreased and the combustion is performed, and inthis mode, the temperature inside the cylinder increased comparativelyhigh in the spark-ignition mode is decreased and the temperature of theexhaust gas discharged after the combustion is also decreased. When theignition is shifted to the compression-ignition mode and the exhaust gasis introduced into the cylinder after performing the transition mode,the exhaust gas with comparatively low temperature is introduced intothe cylinder. As a result, in the compression-ignition mode,pre-ignition caused by the mixture gas inside the cylinder can beavoided, and the mixture gas can be compressed to ignite at a suitabletiming. Thus, when switching from the spark-ignition mode to thecompression-ignition mode, by intervening the transition modetherebetween, a rapid increase in pressure inside the cylinder can beavoided, and as a result, an increase in combustion noises can beavoided as well. Here, the transition mode may be performed over oneengine cycle or continuously over a plurality of engine cycles until thetemperature inside the cylinder and the temperature of the exhaust gasare sufficiently decreased.

The controller may decrease the temperature inside the cylinder andperform the compression-ignition combustion by injecting the fuel withthe fuel injection valve from an intake stroke to an early stage ofcompression stroke in the compression-ignition mode, and at least aftera middle stage of the compression stroke in the transition mode.

Here, “the early stage of the compression stroke” and “the middle stageof the compression stroke” may be defined as the early stage and themiddle stage when the compression stroke is divided into three stages ofan early stage, a middle stage, and a late stage. In this case, “afterthe middle stage of the compression stroke” includes the middle and latestages and expansion stroke.

In the compression-ignition mode, for example, the fuel is injected intothe cylinder from the intake stroke to the early stage of thecompression stroke. By this, comparatively homogeneous mixture gas canbe formed and the compression-ignition combustion can be performedsuitably; however, since the reactable time length of the mixture gasbecomes longer, in a case where the temperature inside the cylinderbecomes excessively high, pre-ignition may occur.

Whereas, in the transition mode, the fuel is injected into the cylinderat least after the middle stage of the compression stroke. In otherwords, the fuel is injected into the cylinder after the intake air isintroduced into the cylinder and the compression stroke is substantiallystarted. Thus, the temperature inside the cylinder is decreased by alatent heat of vaporization of the fuel. Moreover, since a fuelinjection timing is relatively late and the reactable time length of themixture gas becomes short, pre-ignition does not easily occur. Thus, inthe transition mode, even in an environment which is immediately afterthe spark-ignition mode and where the temperature inside the cylindermay be comparatively high, the compression-ignition combustion isperformed suitably, a temperature of combusting gas is decreased by thecompression-ignition combustion with high thermal efficiency, and as aresult, the temperature of the exhaust gas can be decreased.

As a result of performing such a transition mode, the ignition mode canbe shifted to the compression-ignition mode after the temperature insidethe cylinder and the temperature of the exhaust gas are sufficientlydecreased, and in the compression-ignition mode, pre-ignition caused bythe mixture gas inside the cylinder can be avoided. In other words, whenswitching from the spark-ignition mode to the compression-ignition mode,a rapid increase in pressure inside the cylinder can be avoided, and asa result, an increase in combustion noises can be avoided.

Note that, in the transition mode, the fuel may be injected into thecylinder after the middle stage of the compression stroke (i.e., lumpinjection), or divided injections may be performed by injecting the fuelfrom the intake stroke to the early stage of the compression stroke andinjecting the fuel into the cylinder after the middle stage of thecompression stroke. In the case of performing the divided injections,the amount of fuel injected after the middle stage of the compressionstroke is preferred to be relatively large in order to exert asufficient reduction in temperature inside the cylinder.

The controller may increase an amount of fresh air introduced into thecylinder in the transition mode.

The increase in fresh air may be performed by any one of an openingadjustment of a throttle valve and a change in the open/close operationsof the intake and exhaust valves.

By increasing the amount of fresh air introduced into the cylinder inthe transition mode, a sufficient amount of fresh air with a relativelylow temperature is introduced into the cylinder, which becomesadvantageous in decreasing the temperature inside the cylinder. With thecombination of the increased fresh air amount and the fuel injectionafter the middle stage of the compression stroke, an increase incombustion noises can more effectively be avoided when switching fromthe spark-ignition mode to the compression-ignition mode.

In the transition mode, the controller may increase an amount of thefuel injected into the cylinder, corresponding to the increase in thefresh air amount.

By increasing the amount of the fuel introduced into the cylinder, anamount by which the temperature is decreased by the latent heat ofvaporization can be increased. As a result, the increased fuel amountbecomes advantageous in decreasing the temperature inside the cylinder.Moreover, even though the fuel amount is increased, since the amount ofthe fresh air introduced into the cylinder is also increased,degradation of the exhaust emission performance is avoided.

In the transition mode, the controller may perform the spark-ignitioncombustion and decrease a temperature of combusting gas by advancing anignition timing of the ignition plug by at least a minimum advance forbest torque.

With this configuration, in the transition mode, differently from thedescription above, the spark-ignition combustion is performed, and theignition timing thereof is advanced by at least the minimum advance forbest torque (MBT). By this, the temperature of the combusting gas can bedecreased while hardly reducing the generated torque. As a result,similarly to the description above, when the ignition mode is shifted tothe compression-ignition mode after performing the transition mode andthe exhaust gas is introduced into the cylinder, the exhaust gas with acomparatively low temperature is introduced into the cylinder. Thus, inthe compression-ignition mode, pre-ignition caused by the mixture gasinside the cylinder and an increase in combustion noises can be avoided.

The control device may also include an external exhaust gasrecirculation (EGR) adjuster for circulating exhaust gas discharged fromthe cylinder back to the intake side. When switching from thespark-ignition mode where the external EGR adjuster is operated tocirculate the exhaust gas into the cylinder to the compression-ignitionmode, in the transition mode, the controller may perform thespark-ignition combustion and advance the ignition timing of theignition plug by at least the minimum advance for best torque.

The control device may also include an external EGR adjuster forcirculating exhaust gas discharged from the cylinder back to the intakeside. The controller may perform the compression-ignition combustion inthe transition mode when switching from the spark-ignition mode wherethe external EGR adjuster is not operated, to the compression-ignitionmode.

When the engine load is at a predetermined high load, the controller maycontrol the ignition mode to the spark-ignition mode where the externalEGR adjuster is operated, and when the engine load is at a predeterminedlow load, the controller may control the ignition mode to thespark-ignition mode where the external EGR adjuster is not operated.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram illustrating a configuration of aspark-ignition engine.

FIG. 2 is a block diagram relating to a control of the spark-ignitionengine.

FIG. 3 is a cross-sectional view illustrating a combustion chamber in anenlarged manner.

FIG. 4 is a chart exemplifying an operating range of the engine.

FIG. 5A illustrates one example of a fuel injection timing in a casewhere an intake stroke injection is performed in a compression-ignition(CI) mode and a heat release rate of CI combustion accompanied by theintake stroke injection, FIG. 5B illustrates one example of a fuelinjection timing in a case where a high pressure retarded injection isperformed in the CI mode and a heat release rate of the CI combustionaccompanied by the high pressure retarded injection, FIG. 5C illustratesone example of a fuel injection timing and an ignition timing in a casewhere the high pressure retarded injection is performed in aspark-ignition (SI) mode and a heat release rate of the SI combustionaccompanied by the high pressure retarded injection, and FIG. 5Dillustrates one example of a fuel injection timing and an ignitiontiming in a case where split injections of the intake stroke injectionand the high pressure retarded injection are performed in the SI modeand a heat release rate of the SI combustion accompanied by the splitinjections.

FIG. 6 is a composite chart comparing between a state of the SIcombustion caused by the high pressure retarded injection and a state ofthe conventional SI combustion.

FIG. 7 is a time chart for describing a transition control performedwhen switching from a part of the SI mode where an external EGR gas isnot introduced, to the CI mode.

FIG. 8 is a time chart for describing a transition control performedwhen switching from a part of the SI mode where the external EGR gas isintroduced to the CI mode.

FIG. 9 is a flowchart of the transition control when switching from theSI mode to the CI mode.

FIG. 10 is a time chart for describing a transition control differentfrom FIG. 7, performed when switching from the part of the SI mode wherethe external EGR gas is not introduced, to the CI mode.

FIG. 11 is a time chart for describing a transition control differentfrom FIG. 8, performed when switching from the part of the SI mode wherethe external EGR gas is introduced, to the CI mode.

DETAILED DESCRIPTION OF EMBODIMENT

Hereinafter, a spark-ignition engine according to one embodiment of thepresent invention is described in detail with reference to the appendeddrawings. The following description of the preferred embodiment is anillustration. FIGS. 1 and 2 illustrate a schematic configuration of anengine 1 (engine body) of this embodiment. The engine 1 is aspark-ignition gasoline engine that is equipped in a vehicle andsupplied with fuel containing at least gasoline. The engine 1 includes acylinder block 11 provided with a plurality of cylinders 18 (note that,although only one cylinder is illustrated in FIG. 1, four cylinders arelinearly provided in this embodiment), a cylinder head 12 disposed onthe cylinder block 11, and an oil pan 13 disposed below the cylinderblock 11, where a lubricant is stored. Inside the cylinders 18,reciprocatable pistons 14 coupled to a crankshaft 15 via respectiveconnecting rods 142, are fitted. As illustrated in FIG. 3 in an enlargedmanner, a cavity 141 having a reentrant shape, such as the shapegenerally used in a diesel engine, is formed on a top face of eachpiston 14. When the piston 14 is at a position near a compression topdead center (CTDC), the cavity 141 faces toward an injector 67 describedlater. The cylinder head 12, the cylinders 18, and the pistons 14 eachformed with the cavity 141 partition combustion chambers 19. Note thatthe shape of the combustion chamber 19 is not limited to the shape inthe drawings. For example, the shape of the cavity 141, the shape of thetop face of the piston 14, and the shape of a ceiling part of thecombustion chamber 19 may suitably be changed.

A geometric compression ratio of the engine 1 is set comparatively highas 15:1 or above so as to improve a theoretical thermal efficiency andstabilize compression-ignition combustion (described later). Note that,the geometric compression ratio may suitably be set within a range of15:1 to about 20:1. For example, the geometric compression ratio of theengine 1 of this embodiment is 18:1.

In the cylinder head 12, each of the cylinders 18 is formed with anintake port 16 and an exhaust port 17, and disposed with an intake valve21 for opening and closing the intake port 16 on the combustion chamber19 side and an exhaust valve 22 for opening and closing the exhaust port17 on the combustion chamber 19 side.

In a valve train system of the engine 1 for operating the intake andexhaust valves 21 and 22, for example, a hydraulically-actuated variablevalve mechanism 71 (see FIG. 2, hereinafter, may be referred to as theVVL (Variable Valve Lift)) for switching an operation mode of theexhaust valve 22 between a normal mode and a special mode is provided onan exhaust side. The VVL 71 (the detailed configuration is notillustrated) on the exhaust side (exhaust VVL) includes two kinds ofcams with different cam profiles in which a first cam has one cam noseand a second cam has two cam noses; and a lost motion mechanism forselectively transmitting an operating state of either one of the firstand second cams to the exhaust valve. While the lost motion mechanismtransmits the operating state of the first cam to the exhaust valve 22,the exhaust valve 22 operates in the normal mode where it opens onlyonce during exhaust stroke. On the other hand, while the lost motionmechanism transmits the operating state of the second cam to the exhaustvalve 22, the exhaust valve 22 operates in the special mode, which is aso-called exhaust open-twice control, where it opens once during theexhaust stroke and once more during an intake stroke (see FIG. 7, etc.).The normal and special modes of the exhaust VVL 71 are switchedtherebetween according to an operating state of the engine.Specifically, the special mode is utilized for a control related to aninternal EGR. Hereinafter, the processing of operating the exhaust VVL71 in the normal mode where the exhaust open-twice control is notperformed may be referred to as “turning the VVL 71 off,” and theprocessing of operating the exhaust VVL 71 in the special mode where theexhaust open-twice control is performed may be referred to as “turningthe VVL 71 on.”

Here, the exhaust open-twice control includes, in addition to a liftproperty in which the exhaust valve 22 is substantially closed on theexhaust stroke and then opened again on the intake stroke (i.e., thelift property in which two noses of a lift curve of the exhaust valve 22align in a progressing direction of a crank angle), a lift property inwhich the exhaust valve 22 lifted on the exhaust stroke once maintains apredetermined opening until the intake stroke without being closed(i.e., the lift property in which although the nose of the lift curve ofthe exhaust valve 22 is substantially one, the foot part of the noseextends in the progressing direction of the crank angle). Note that inenabling the switching between the normal mode and the special mode, anelectromagnetic valve train system which operates the exhaust valve 22with an electromagnetic actuator may be adopted.

Moreover, the execution of the internal EGR is not limited to beachieved by the exhaust open-twice control only. For example, theinternal EGR control may be performed by an intake open-twice control inwhich the intake valve 21 opens twice. Similarly to the exhaustopen-twice control, the intake open-twice control includes, in additionto a lift property in which the intake valve 21 is substantially closedon the exhaust stroke and then opened again on the intake stroke (i.e.,the lift property in which two noses of the lift curve of the intakevalve 21 align in a progressing direction of the crank angle), a liftproperty in which the intake valve 21 lifted on the exhaust stroke oncemaintains a predetermined opening until the intake stroke without beingclosed (i.e., the lift property in which although the nose of the liftcurve of the intake valve 21 is substantially one, the foot part of thenose extends in an opposite direction to the progressing direction ofthe crank angle). Moreover, an internal EGR control may be performed inwhich burned gas remains inside the cylinder 18 by providing a negativeoverlap period in which both the intake and exhaust valves 21 and 22 areclosed on the exhaust stroke or the intake stroke.

An intake side of the valve train system includes a VVL 73 similarly tothe exhaust side of the valve train system provided with the VVL 71.Note that, the VVL 73 on the intake side (intake VVL), differently fromthe VVL 71 on the exhaust side, includes two kinds of cams withdifferent cam profiles in which a large lift cam relatively increasesthe lift of the intake valve 21 and a small lift cam relatively reducesthe lift of the intake valve 21; and a lost motion mechanism forselectively transmitting an operating state of either one of the largeand small lift cams to the intake valve 21. While the lost motionmechanism transmits the operating state of the large lift cam to theintake valve 21, the intake valve 21 opens with a relatively large lift,and an open period thereof is long. On the other hand, while the lostmotion mechanism transmits the operating state of the small lift cam tothe intake valve 21, the intake valve 21 opens with a relatively smalllift, and the open period thereof is long (see FIG. 7, etc.). Moreover,as illustrated in FIG. 2, a phase variable mechanism 72 (hereinafter,may be referred as the VVT (Variable Valve Timing)) for changing arotational phase of an intake camshaft with respect to the crankshaft 15is provided on the intake side of the valve train system. A knownhydraulic, electromagnetic, or mechanical structure may suitably beadopted for the VVT 72 (a detailed structure is not illustrated). Theopen and close timings and the lift of the intake valve 21 can bechanged by the VVT 72 and the VVL 73, respectively.

For each cylinder 18, the injector 67 for directly injecting the fuelinto the cylinder 18 is attached to the cylinder head 12. As illustratedin an enlarged manner in FIG. 3, a nozzle hole of the injector 67 isarranged in a center portion of the ceiling face of the combustionchamber 19 to be oriented toward the inside of the combustion chamber19. The injector 67 directly injects the fuel into the combustionchamber 19 by an amount according to the operating state of the engine 1at an injection timing set according to the operating state of theengine 1. In this embodiment, the injector 67 (a detailed configurationis not illustrated) is a multi hole injector formed with a plurality ofnozzle holes. Thus, the injector 67 injects the fuel so that the fuelspray spreads radially from the center portion of the combustion chamber19. As indicated by the arrows in FIG. 3, at a timing when the piston 14reaches near the CTDC, the fuel spray injected to spread radially fromthe center portion of the combustion chamber 19 flows along a wallsurface of the cavity 141 formed on the piston top face. Therefore, itmay be said that the cavity 141 is formed to contain therewithin thefuel spray injected at the timing when the piston 14 reaches near theCTDC. The combination of the multi hole injector 67 and the cavity 141is advantageous in, after the fuel is injected, shortening a mixture gasforming period and the combustion period. Note that, the injector 67 isnot limited to the multi hole injector, and may be an outward openingvalve type injector.

A fuel supply path couples between a fuel tank (not illustrated) and theinjectors 67. A fuel supply system 62 for supplying the fuel to each ofthe injectors 67 at a comparatively high fuel pressure and having a fuelpump 63 and a common rail 64 is provided within the fuel supply path.The fuel pump 63 pumps the fuel from the fuel tank to the common rail64, and the common rail 64 can accumulate the pumped fuel at acomparatively high fuel pressure. By opening the nozzle holes of theinjector 67, the fuel accumulated in the common rail 64 is injected fromthe nozzle holes of the injector 67. Here, the fuel pump 63 is a plungertype pump (not illustrated) and is operated by the engine 1. The fuelsupply system 62 including the engine-operated pump enables the supplyof the fuel to the injector 67 at a high fuel pressure of 30 MPa orabove. The fuel pressure may be set to about 120 MPa at the maximum. Asdescribed later, the pressure of the fuel supplied to the injector 67 ischanged according to the operating state of the engine 1. Note that thefuel supply system 62 is not limited to the above configuration.

Further, as illustrated in FIG. 3, an ignition plug 25 for ignitingmixture gas inside the combustion chamber 19 is attached to the cylinderhead 12 for each cylinder 18. In this embodiment, the ignition plug 25is arranged penetrating the cylinder head 12 so as to extend obliquelydownward from the exhaust side of the engine 1. As illustrated in FIG.3, a tip of the ignition plug 25 is oriented toward the inside of thecavity 141 of the piston 14 at the CTDC.

On one side surface of the engine 1, as illustrated in FIG. 1, an intakepassage 30 is connected to communicate with the intake port 16 of eachcylinder 18. On the other side surface of the engine 1, an exhaustpassage 40 is connected to guide out the burned gas (exhaust gas)discharged from each of the combustion chambers 19 of the cylinders 18.

An air cleaner 31 for filtrating intake air is disposed in an upstreamend part of the intake passage 30. A surge tank 33 is disposed near adownstream end of the intake passage 30. A part of the intake passage 30downstream of the surge tank 33 is branched to be independent passagesextending toward the respective cylinders 18, and downstream ends of theindependent passages are connected with the intake ports 16 of thecylinders 18, respectively.

A water-cooled type intercooler/warmer 34 for cooling or heating air anda throttle valve 36 for adjusting an intake air amount to each cylinder18 are disposed between the air cleaner 31 and the surge tank 33 in theintake passage 30. Moreover, an intercooler bypass passage 35 forbypassing the intercooler/warmer 34 is connected to the intake passage30, and an intercooler bypass valve 351 for adjusting an air flow ratepassing through the passage 35 is disposed within the intercooler bypasspassage 35. A ratio of a flow rate within the intercooler bypass passage35 with a flow rate within the intercooler/warmer 34 is adjusted throughcontrolling an opening of the intercooler bypass valve 351, and thus, atemperature of fresh air introduced into the cylinder 18 can beadjusted.

An upstream part of the exhaust passage 40 includes an exhaust manifold.The exhaust manifold has independent passages branched toward therespective cylinders 18 and connected with respective external ends ofthe exhaust ports 17, and a manifold section where the independentpassages merge together. In a part of the exhaust passage 40 on thedownstream side of the exhaust manifold, a direct catalyst 41 and anunderfoot catalyst 42 are connected as an exhaust emission controlsystem for purifying hazardous components contained in the exhaust gas.Each of the direct catalyst 41 and the underfoot catalyst 42 includes acylindrical case and, for example, a three-way catalyst disposed in aflow path within the case.

A part of the intake passage 30 between the surge tank 33 and thethrottle valve 36 is connected with a part of the exhaust passage 40 onthe upstream side of the direct catalyst 41 via an EGR passage 50 forcirculating a part of the exhaust gas back to the intake passage 30. TheEGR passage 50 includes a main passage 51 disposed with an EGR cooler 52for cooling the exhaust gas by an engine coolant, and an EGR coolerbypass passage 53 for bypassing the EGR cooler 52. An EGR valve 511 foradjusting a circulation amount of the exhaust gas to the intake passage30 is disposed within the main passage 51. An EGR cooler bypass valve531 for adjusting a flow rate of the exhaust gas flowing through the EGRcooler bypass passage 53 is disposed within the EGR cooler bypasspassage 53.

The engine 1 with the configuration described as above is controlled bya powertrain control module 10 (hereinafter, may be referred to as thePCM). The PCM 10 is comprised of a microprocessor including a CPU, amemory, a counter timer group, an interface, and paths for connectingthese units. The PCM 10 configures the controller.

As illustrated in FIGS. 1 and 2, detection signals of various kinds ofsensors SW1 to SW16 are inputted to the PCM 10. The various kinds ofsensors include the following sensors: an air flow sensor SW1 fordetecting the flow rate of the fresh air and an intake air temperaturesensor SW2 for detecting the temperature of the fresh air that arearranged on the downstream side of the air cleaner 31; a second intakeair temperature sensor SW3, arranged on the downstream side of theintercooler/warmer 34, for detecting the temperature of the fresh airafter passing through the intercooler/warmer 34; an EGR gas temperaturesensor SW4, arranged near a connecting part of the EGR passage 50 withthe intake passage 30, for detecting a temperature of external EGR gas;an intake port temperature sensor SW5, attached to the intake port 16,for detecting the temperature of the intake air immediately beforeflowing into the cylinder 18; an in-cylinder pressure sensor SW6,attached to the cylinder head 12, for detecting the pressure inside thecylinder 18; an exhaust gas temperature sensor SW7 and an exhaust gaspressure sensor SW8, arranged near a connecting part of the exhaustpassage 40 with the EGR passage 50, for detecting the exhaust gastemperature and pressure, respectively; a linear O₂ sensor SW9, arrangedon the upstream side of the direct catalyst 41, for detecting an oxygenconcentration within the exhaust gas; a lambda O₂ sensor SW10, arrangedbetween the direct catalyst 41 and the underfoot catalyst 42, fordetecting an oxygen concentration within the exhaust gas; a fluidtemperature sensor SW11 for detecting a temperature of the enginecoolant; a crank angle sensor SW12 for detecting a rotational angle ofthe crankshaft 15; an accelerator position sensor SW13 for detecting anaccelerator opening corresponding to an angle of an acceleration pedal(not illustrated) of the vehicle; an intake cam angle sensor SW14 and anexhaust cam angle sensor SW 15; and a fuel pressure sensor SW16,attached to the common rail 64 of the fuel supply system 62, fordetecting the fuel pressure supplied to the injector 67.

By performing various kinds of operations based on these detectionsignals, the PCM 10 determines the state of the engine 1 and further thevehicle, and outputs control signals to the injectors 67, the ignitionplugs 25, the VVT 72 and the intake VVL 73, the exhaust VVL 71, the fuelsupply system 62, and the actuators of the various kinds of valves (thethrottle valve 36, the intercooler bypass valve 351, the EGR valve 511,and the EGR cooler bypass valve 531) according to the determined state.In this manner, the PCM 10 operates the engine 1.

FIG. 4 illustrates one example of an operating range of the engine 1 ina warmed-up state. Within a low engine load range where an engine loadis relatively low, the engine 1 performs compression-ignition combustionby a combustion generated from a compression self-ignition withoutperforming an ignition by the ignition plug 25, so as to improve fuelconsumption and exhaust emission performance. However, with thecompression-ignition combustion, the speed of the combustion becomesexcessively rapid as the engine load increases, causing a problem ofcombustion noises, etc. Therefore, with the engine 1, within a highengine load range where the engine load is relatively high, thecompression-ignition combustion is suspended and is switched to aspark-ignition combustion using the ignition plug 25. As describedabove, the engine 1 is configured to switch a combustion mode accordingto the operating state of the engine 1, particularly according to theload of the engine 1, between a CI (Compression-Ignition) mode where thecompression-ignition combustion is performed and an SI (Spark-Ignition)mode where the spark-ignition combustion is performed. Note that, theboundary of switching the mode is not limited to the example in theillustration. Moreover, as described later, the engine 1 is configuredto switch the mode under various circumstances according to itsoperating state, other than the level of the engine load.

The CI mode is divided into two ranges according to the level of theengine load. Specifically, within a range (1) where the engine load islow to medium within the CI mode, hot EGR gas with relatively hightemperature is introduced into the cylinder 18 to improve ignitabilityand stability of the compression-ignition combustion. This is achievedby turning the exhaust VVL 71 on and performing the exhaust open-twicecontrol of opening the exhaust valve 22 during the intake stroke. Theintroduction of the hot EGR gas increases the in-cylinder temperature atthe end of the compression stroke, and is advantageous in improving theignitability and the stability of the compression-ignition combustionwithin the range (1). Moreover, within the range (1), as illustrated inFIG. 5A, the injector 67 injects the fuel into the cylinder 18 at leastin a period from the intake stroke to the middle stage of thecompression stroke, and thus homogeneous mixture gas is formed. Withinthe range (1), the air-fuel ratio of the mixture gas is basically set tothe theoretical air-fuel ratio (A/F=14.7±0.5, an air excess ratio λ≈1).Note that, as indicated by the dashed-line in FIG. 4, within a part ofthe range (1) where the engine load and the engine speed are relativelylow, the air-fuel ratio of the mixture gas is set leaner than thetheoretical air-fuel ratio.

Thus, within the range (1), the mixture gas inside the combustionchamber 19 is compressed to self-ignite near the CTDC as illustrated inFIG. 5A.

In the CI mode, within a range (2) where the engine load is higher thanthe range (1), the air-fuel ratio of the mixture gas is set to thetheoretical air-fuel ratio (λ≈1). By setting to the theoretical air-fuelratio, the three-way catalyst can be used and, as described later, theair-fuel ratio of the mixture gas becomes the theoretical air-fuel ratioalso in the SI mode. Thus, the control performed when switching betweenthe SI mode and the CI mode is simplified, and moreover, it contributesto expanding the range of the CI mode to the higher engine load side.

Moreover, within the range (2), since the in-cylinder temperaturenaturally increases according to the increase of the engine load, thehot EGR gas amount is reduced to avoid pre-ignition. This reduction isachieved by adjusting the internal EGR gas amount introduced into thecylinder 18. Moreover, by adjusting the amount of external EGR gasbypassing the EGR cooler 52, the amount of hot EGR gas may be adjusted.

Furthermore, within the range (2), cooled EGR gas with a relatively lowtemperature is introduced into the cylinder 18. Thus, by introducing thehot EGR gas with a high temperature and the cooled EGR gas with a lowtemperature into the cylinder 18 at a suitable ratio, the in-cylindertemperature at the end of the compression stroke is adjusted suitably,rapid combustion is avoided while securing the ignitability of thecompression-ignition, and the compression-ignition combustion isstabilized.

Thus, within the range (2) including the switching boundary between theCI mode and the SI mode, although the in-cylinder temperature isdecreased, the in-cylinder temperature may further increase at the endof the compression stroke. If the fuel is injected into the cylinder 18in the period from the intake stroke to the middle stage of thecompression stroke similarly to the range (1), it may cause abnormalcombustion (e.g., pre-ignition). On the other hand, if a large amount ofcooled EGR gas with a low temperature is introduced to decrease thein-cylinder temperature at the end of the compression stroke, then theignitability of the compression-ignition will degrade. In other words,since the compression-ignition combustion cannot be performed stablyonly by controlling the in-cylinder temperature, within the range (2),by adjusting the fuel injection mode in addition to the in-cylindertemperature control, the compression-ignition combustion can bestabilized while avoiding abnormal combustion (e.g., pre-ignition).Specifically, in this fuel injection mode, as illustrated in FIG. 5B,the fuel is injected into the cylinder 18 at least in a period from thelate stage of the compression stroke and the early stage of expansionstroke (hereinafter, referred to as the retard period) at asignificantly higher fuel pressure compared to the conventional mode.Hereinafter, this characteristic fuel injection mode is referred to asthe “high pressure retarded injection” or simply “retarded injection.”By the high pressure retarded injection, the compression-ignitioncombustion can be stabilized while avoiding the abnormal combustionwithin the range (2). The details of the high pressure retardedinjection will be described later.

While the CI mode has the two divided ranges according to the level ofthe engine load, the SI mode is divided into two ranges (3) and (4)according to the level of engine speed. When the operating range of theengine 1 is divided into two high and low speed ranges, in FIG. 4, therange (3) corresponds to the low speed range and a lower load part ofthe high speed range, and the range (4) corresponds to a higher loadpart of the high speed range. Note that, the boundary between the ranges(3) and (4) is not limited to the illustration.

In each of the ranges (3) and (4), the mixture gas is set to thetheoretical air-fuel ratio (λ≈1) similarly to the range (2). Therefore,the air-fuel ratio of the mixture gas is fixed at the theoreticalair-fuel ratio (λ≈1) over the boundary between the CI mode and the SImode. Moreover, in the SI mode (i.e., the ranges (3) and (4)), thethrottle valve 36 is basically fully opened and the fresh air amount andthe external EGR gas amount introduced into the cylinder 18 are adjustedby controlling an opening of the EGR valve 511. Note that, even withinthe range of the SI mode, within a part of the range where the engineload is relatively low, the throttle valve 36 may be throttled. Theadjustment of the ratio of gas introduced into the cylinder 18 reduces apumping loss, and by introducing a large amount of EGR gas into thecylinder 18, the temperature of the spark-ignition combustion issuppressed low and cooling loss can be reduced. Within the ranges of theSI mode, the external EGR gas cooled mainly by passing through the EGRcooler 52 is introduced into the cylinder 18. This becomes advantageousin avoiding abnormal combustion as well as suppressing generation of rawNOx. Note that, within a full engine load range, the EGR valve 511 isfully closed to cancel the external EGR.

Note that, within the range of the SI mode, the fresh air amountintroduced into the cylinder 18 may be adjusted to set the air-fuelratio to the theoretical air-fuel ratio (1) by controlling the openingof the throttle valve 36 according to the fuel injection amount whilesuspending the introduction of the EGR gas.

The geometric compression ratio of the engine 1 is, as described above,set to 15:1 or above (e.g., 18:1). Since a high compression ratioincreases the in-cylinder temperature and the in-cylinder pressure atthe end of the compression stroke, it is advantageous in stabilizing thecompression-ignition combustion in the CI mode, especially within a lowengine load part of the range of the CI mode (e.g., the range (1)).Whereas, in the SI mode corresponding to the high engine load range,such a high compression ratio causes a problem in engine 1 in thatabnormal combustion (e.g., pre-ignition and knocking) easily occurs.

Thus, with the engine 1, the high pressure retarded injection isperformed within the ranges (3) and (4) of the SI mode to avoid abnormalcombustion. Specifically, within the range (3), at a high fuel pressureof 30 MPa or above, as illustrated in FIG. 5C, only the high pressureretarded injection of injecting the fuel into the cylinder 18 isperformed in the retard period from the late stage of the compressionstroke to the early stage of the expansion stroke. On the other hand,within the range (4), as illustrated in FIG. 5D, part of the fuelinjected is injected into the cylinder 18 in an intake stroke periodwhere the intake valve 21 opens, and the rest of the fuel injected isinjected into the cylinder 18 in the retard period. In other words,within the range (4), a split injection of the fuel is performed. Here,the intake stroke period where the intake valve 21 opens is a perioddefined based on open and close timings of the intake valve, and not aperiod defined based on the piston position. Here, the end of the intakestroke may vary with respect to the timing at which the piston reachesan intake bottom dead center (IBDC) depending on the close timing of theintake valve 21 which is changed by the VVL 73 and the VVT 72.

Next, the high pressure retarded injection in the SI mode is describedwith reference to FIG. 6, which shows a composite chart comparingdifferences in a heat release rate (upper chart) and an extent ofreaction of unburned mixture gas (lower chart) between an SI combustioncaused by the high pressure retarded injection described above (solidline) and the conventional SI combustion in which the fuel injection isperformed during the intake stroke (dashed line). The lateral axis inFIG. 6 indicates the crank angle. The comparison is performed under acondition that the operating state of the engine 1 is within the lowerengine speed range with a high engine load (i.e., the range (3)), and afuel amount injected is the same between the SI combustion caused by thehigh pressure retarded injection and the conventional SI combustion.

First, for the conventional SI combustion, a predetermined amount offuel is injected into the cylinder 18 during the intake stroke (dashedline in the upper chart). After the fuel injection, comparativelyhomogeneous mixture gas is formed inside the cylinder 18 before thepiston 14 reaches the CTDC. Further, in this case, the ignition isperformed at a predetermined timing indicated by the first white circleafter the CTDC to start the combustion. After the combustion starts, asindicated by the dashed line in the upper chart of FIG. 6, thecombustion ends after progressing through a peak in the heat releaserate. A period from the start of the fuel injection until the end of thecombustion corresponds to a reactable time length of unburned mixturegas (hereinafter, may simply be referred to as the reactable timelength) and, as indicated by the dashed line in the lower chart of FIG.6, the reaction of the unburned mixture gas gradually progresses withinthe reactable time length. The dotted line in the lower chart indicatesan ignition threshold (i.e., a reactivity of the unburned mixture gasbeing ignited). The conventional SI combustion is performed within thelow engine speed range and it has an extremely long reactable timelength. The reaction of the unburned mixture gas keeps progressing forthe reactable time length, and therefore, the reactivity of the unburnedmixture gas exceeds the ignition threshold around the ignition timing,causing abnormal combustion (e.g., pre-ignition and knocking).

On the other hand, the high pressure retarded injection aims to avoidabnormal combustion by shortening the reactable time length. Asillustrated in FIG. 6, the reactable time length in this case is a totaltime length of a period where the injector 67 injects the fuel ((1) aninjection period), a period from the end of the injection untilcombustible mixture gas is formed around the ignition plug 25 ((2) amixture gas forming period), and a period from the start of thecombustion started by the ignition until the combustion ends ((3) acombustion period), in other words, (1)+(2)+(3). The high pressureretarded injection shortens each of the injection period, the mixturegas forming period, and the combustion period, and thereby, shortens thereactable time length. The methods of shortening the periods areexplained sequentially below.

First, a high fuel pressure relatively increases the fuel amountinjected from the injector 67 per unit time. Therefore, in a case wherethe fuel injection amount is fixed, a relationship between the fuelpressure and the injection period of the fuel substantially becomes asfollows: the injection period extends as the fuel pressure decreases,and the injection period contracts as the fuel pressure increases.Therefore, the high pressure retarded injection in which the fuelpressure is set significantly higher than the conventional pressureshortens the injection period.

Further, the high fuel pressure is advantageous in atomizing the fuelspray injected into the cylinder 18 and further extends a spreadingdistance of the fuel spray. Therefore, a relationship between the fuelpressure and a fuel vaporization time length substantially becomes asfollows: the fuel vaporization time length extends as the fuel pressuredecreases, and the fuel vaporization time length contracts as the fuelpressure increases. Further, a relationship between the fuel pressureand a time length for the fuel spray to reach around the ignition plug25 (the fuel spray reaching time length) substantially becomes asfollows: the fuel spray reaching time length extends as the fuelpressure decreases, and the fuel spray reaching time length contracts asthe fuel pressure increases. The mixture gas forming period correspondsto a total time length of the fuel vaporization time length and the fuelspray reaching time length required to reach around the ignition plug25; therefore, the mixture gas forming period contracts as the fuelpressure increases. Therefore, the high pressure retarded injection inwhich the fuel pressure is set significantly higher than theconventional pressure shortens the fuel vaporization time length and thefuel spray reaching time length required to reach around the ignitionplug 25 and, as a result, shortens the mixture gas forming period. Onthe other hand, as indicated by the white circles of the chart in FIG.6, with the conventional intake stroke injection with the low fuelpressure, the mixture gas forming period is significantly longer. Notethat, in the SI mode, the combination of the multi hole injector 67 andthe cavity 141 shortens the time length from the end of the fuelinjection until the fuel spray reaches around the ignition plug 25 and,as a result, becomes advantageous in shortening the mixture gas formingperiod.

As described above, shortening the injection period and the mixture gasforming period enables retarding the injection timing of the fuel, moreprecisely, retarding the injection start timing, to a comparatively latetiming. Therefore, with the high pressure retarded injection, asillustrated in the upper chart of FIG. 6, the fuel injection isperformed within the retard period from the late stage of thecompression stroke to the early stage of the expansion stroke. Althoughturbulence of flow inside the cylinder becomes stronger and a turbulencekinetic energy inside the cylinder 18 increases due to injecting thefuel into the cylinder 18 at the high fuel pressure, the high turbulencekinetic energy is advantageous in shortening the combustion period, incombination with retarding the fuel injection timing to thecomparatively late timing.

In other words, in a case where the fuel injection is performed withinthe retard period, a relationship between the fuel pressure and theturbulence kinetic energy within the combustion period substantiallybecomes as follows: the turbulence kinetic energy decreases as the fuelpressure decreases, and the turbulence kinetic energy increases as thefuel pressure increases. Here, even if the fuel is injected into thecylinder 18 at the high fuel pressure, in the case where the injectiontiming is on the intake stroke, due to the time length until theignition timing being long and the inside of the cylinder 18 beingcompressed on the compression stroke after the intake stroke, theturbulence inside the cylinder 18 is subsided. As a result, in the casewhere the fuel injection is performed during the intake stroke, theturbulence kinetic energy within the combustion period becomescomparatively low regardless of the fuel pressure.

A relationship between the turbulence kinetic energy within thecombustion period and the combustion period substantially becomes asfollows: the combustion period extends as the turbulence kinetic energydecreases and the combustion period contracts as the turbulence kineticenergy increases. Therefore, a relationship between the fuel pressureand the combustion period becomes as follows: the combustion periodextends as the fuel pressure decreases and the combustion periodcontracts as the fuel pressure increases. In other words, the highpressure retarded injection shortens the combustion period. On the otherhand, with the conventional intake stroke injection with the low fuelpressure, the combustion period extends. Note that, the multi holeinjector 67 is advantageous in increasing the turbulence kinetic energyinside the cylinder 18 and shortening the combustion period. Moreover,it is also advantageous in shortening the combustion period to keep thefuel spray contained within the cavity 141 by the combination of themulti hole injector 67 and the cavity 141.

As above, the high pressure retarded injection shortens each of theinjection period, the mixture gas forming period, and the combustionperiod, and as a result, as illustrated in FIG. 6, the high pressureretarded injection can significantly shorten the reactable time lengthof the unburned mixture gas from a fuel injection start timing SOI to acombustion end timing Oend compared to the conventional case where thefuel injection is performed during the intake stroke. As a result ofshortening the reactable time length, as illustrated in the upper chartof FIG. 6, while the extent of reaction of the unburned mixture gas atthe end of the combustion exceeds the ignition threshold and abnormalcombustion occurs with the conventional intake stroke injection with thelow fuel pressure as indicated by the white circle, with the highpressure retarded injection, as indicated by the black circle, theprogression of the reaction of the unburned mixture gas at the end ofthe combustion is suppressed and abnormal combustion can be avoided.Note that, the ignition timings for the cases indicated by the white andblack circles in the upper chart of FIG. 6 are set at the same timing.

By setting the fuel pressure to, for example, 30 MPa or above, thecombustion period can effectively be shortened. Moreover, the fuelpressure of 30 MPa or above can also effectively shorten the injectionperiod and the mixture gas forming period. Note that the fuel pressuremay be suitably set according to a type of fuel used which at leastcontains gasoline. The upper limit value of the fuel pressure may be 120MPa, etc.

The high pressure retarded injection avoids the occurrence of abnormalcombustion in the SI mode by adjusting the mode of the fuel injectioninto the cylinder 18. Alternatively to such high pressure retardedinjection, a conventionally known method for avoiding abnormalcombustion is by retarding the ignition timing. The retarded ignitiontiming suppresses the increase in the temperature and the pressure ofthe unburned mixture gas and, thereby, suppresses the progression of thereaction of the unburned mixture gas. However, while the retardedignition timing causes degradation of the thermal efficiency andreduction of the torque, in the case of performing the high pressureretarded injection, since abnormal combustion is avoided by adjustingthe mode of the fuel injection, the ignition timing can be advanced, andthus, the thermal efficiency can be improved and the torque can beincreased. In other words, the high pressure retarded injection can, notonly avoid abnormal combustion, but also enable advancing the ignitiontiming accordingly, and thereby, is advantageous in improving fuelconsumption.

As described above, the high pressure retarded injection in the SI modecan shorten each of the injection period, the mixture gas formingperiod, and the combustion period, while the high pressure retardedinjection performed within the range (2) of the CI mode can shorten theinjection period and the mixture gas forming period. In other words, byinjecting the fuel at the high fuel pressure into the cylinder 18 toincrease the turbulence inside the cylinder 18, the atomized fuel ismore finely mixed, and even when the fuel injection is performed at thelate timing near the CTDC, the comparatively homogeneous mixture gas canswiftly be formed.

With the high pressure retarded injection in the CI mode, by injectingthe fuel at the late timing near the CTDC within the comparatively highengine load range, substantially homogeneous mixture gas is swiftlyformed as described above while preventing pre-ignition in, for example,a compression stroke period. Therefore, after the CTDC, thecompression-ignition can surely be performed. Further, by performing thecompression-ignition combustion in an expansion stroke period where thepressure inside the cylinder 18 gradually decreases due to motoring, thecombustion subsides, and an excessive increase of the pressure (dP/dθ)inside the cylinder 18 due to the compression-ignition combustion can beavoided. Thus, a restriction due to noise, vibration, and harshness(NVH) is lifted and, as a result, of the CI mode applicable rangeextends to the high load range side.

Back to the SI mode, as described above, the high pressure retardedinjection in the SI mode shortens the reactable time length of theunburned mixture gas by performing the fuel injection in the retardperiod; however, although the shortening of the reactable time length isadvantageous within the low engine speed range where the engine speed iscomparatively low because the actual reactable time length against thecrank angle change is long, within the high engine speed range where theengine speed is comparatively high, since the actual reactable timelength against the crank angle change is short, it is less advantageous.On the other hand, with the retarded injection, since the fuel injectiontiming is set to near the CTDC, on the compression stroke, thein-cylinder gas that does not include the fuel, in other words, air at ahigh specific heat ratio is compressed. As a result, within the highengine speed range, the in-cylinder temperature at the end of thecompression stroke becomes high, and this increased in-cylindertemperature at the end of the compression stroke may cause knocking.Therefore, when only performing the retarded injection within the range(4) where the engine load and the engine speed are high and the fuelamount injected increases, there may be a case where it is required toretard the ignition timing to avoid knocking.

Therefore, within the range (4) where the engine speed is relativelyhigh and the engine load is high in the SI mode as illustrated in FIG.4, part of the fuel injected is injected into the cylinder 18 in theintake stroke period and the rest of the fuel injected is injected intothe cylinder 18 in the retard period, as illustrated in FIG. 5D. Withthe intake stroke injection, the specific heat ratio of the in-cylindergas on the compression stroke (i.e., the mixture gas including the fuel)may be reduced to suppress the in-cylinder temperature at the end of thecompression stroke low. By decreasing the in-cylinder temperature at theend of the compression stroke as above, knocking can be suppressed and,therefore, the ignition timing can be advanced.

Moreover, by performing the high pressure retarded injection, asdescribed above, the turbulence inside the cylinder 18 (in thecombustion chamber 19) near the CTDC becomes strong, and the combustionperiod becomes shorter. This shorter combustion period is alsoadvantageous in suppressing knocking, and the ignition timing canfurther be advanced. Thus, within the range (4), by performing the splitinjection including the intake stroke injection and the high pressureretarded injection, the thermal efficiency can be improved whileavoiding abnormal combustion.

Note that instead of performing the high pressure retarded injection, amulti-point ignition system may be adopted to shorten the combustionperiod within the range (4). Specifically, a plurality of ignition plugsare arranged to be oriented toward the inside of the combustion chamber19, and within the range (4), the intake stroke injection is performedand each of the plurality of ignition plugs is controlled to perform amulti-point ignition. In this case, since a flame spreads from each ofthe plurality of fire sources inside the combustion chamber 19, theflame spreads rapidly and the combustion period becomes shorter. As aresult, the combustion period is shortened similarly to when adoptingthe high pressure retarded injection, and this shortened combustionperiod is advantageous in improving the thermal efficiency.

(Control when Switching from SI Mode to CI Mode)

Since spark-ignition combustion has a low thermal efficiency compared tocompression-ignition combustion, the combusting gas temperature isrelatively high with the spark-ignition combustion. On the other hand,in the CI mode where the compression-ignition combustion is performed,since the ignitability of the compression-ignition is secured asdescribed above, at least the internal EGR gas is introduced into thecylinder 18 to increase the temperature inside the cylinder 18.

Immediately after the SI mode where the combusting gas temperature isrelatively high is switched to the CI mode, since the state inside thecylinder 18 is a high temperature environment and the exhaust gas withhigh temperature caused by the spark-ignition combustion is introducedinto the cylinder 18. Thus, the compression-ignition combustion isperformed while the temperature inside the cylinder 18 is high. In thiscase, if the fuel is injected into the cylinder 18 at a comparativelyearly timing (e.g., during the intake stroke), the pre-ignition iscaused in the compression stroke period, and the pressure increase rate(dP/dθ) inside the cylinder 18 may become significantly high and causeloud combustion noises.

Therefore, with the engine 1, a transition control for avoiding thepre-ignition when switching from the SI mode to the CI mode and avoidingthe increase in combustion noises is performed.

Here, in the operation map in the warmed-up state illustrated in FIG. 4,the switch from the SI mode to the CI mode may correspond to shiftingfrom the high engine load range where the load of the engine 1corresponds to the SI mode to the low engine load range where the loadof the engine 1 corresponds to the CI mode. In other words, due to thereduction of the load of the engine 1, the SI mode is switched to the CImode. Note that, near the boundary between the SI and CI modes, the SImode may be switched to the CI mode in the state where the load of theengine 1 is stable.

Moreover, in a cold-start or a warming-up state where the temperature ofthe engine 1 is below a predetermined temperature, since thecompression-ignition combustion is not stable, the CI mode is notperformed (not illustrated), and the SI mode is performed instead in theentire operating range of the engine 1. On the other hand, asillustrated in FIG. 4, the CI mode is performed in the warmed-up statewhere the temperature of the engine 1 is above the predeterminedtemperature. Therefore, the SI mode may be switched to the CI mode whilethe engine load is stable according to the temperature of the engine 1increasing to the warmed-up state.

Moreover, in view of the combustion stability, since the engine 1 is inthe SI mode in an idle state, when shifting from the idle state to thelow engine load range where the CI mode is performed, the switch isperformed from the low engine load range of the SI mode to the lowengine load range of the CI mode. Additionally, the engine 1 isconfigured to execute a fuel cut while the vehicle decelerates. Sincethe in-cylinder temperature decreases during the fuel cut, immediatelyafter resuming from the fuel cut, there are cases where thecompression-ignition combustion cannot be performed. Therefore, with theengine 1 of this embodiment, immediately after resuming from the fuelcut, the mode is set to the SI mode even within the range of the CI modeto secure the combustion stability, and when the in-cylinder temperatureincreases to shift to the normal CI mode thereafter, the switch isperformed from the low engine load range of the SI mode to the lowengine load range of the CI mode. As described above, the switch fromthe low engine load range of the SI mode to the low engine load range ofthe CI mode, in other words, the switch from the SI mode to the CI modewithout a substantial change in the load of the engine 1, is performedunder various circumstances.

FIG. 7 is a time chart for a transition control performed when switchingfrom the part of the SI mode where the external EGR gas is notintroduced, to the CI mode. Specifically, FIG. 7 illustrates one exampleof a change of the fuel injection timing and the spark-ignition timing,a change of the in-cylinder pressure, a change of the open state of theintake and exhaust vales, a change of the opening of the throttle valve,and a change of a gas state inside the cylinder, when switching from theSI mode to the CI mode. Here, the crank angle (i.e., time) progresses inthe direction from left to right of the time chart in FIG. 7. Note thatthe change of the fuel injection timing, the spark-ignition timing, andthe in-cylinder pressure illustrated in FIG. 7 are merely examples fordescribing this embodiment, and it is not limited to the illustratedtimings (similar to FIG. 8, etc.). Here, the part of the SI mode wherethe external EGR gas is not introduced corresponds, as an example, to asituation in which the SI mode is performed while the engine load islow. The time chart in FIG. 7 corresponds to a situation of switchingfrom the low engine load range of the SI mode to the low engine loadrange of the CI mode.

First, in the first cycle corresponding to the leftmost part in FIG. 7,the engine is operated in the SI mode, and here, the fuel is injected inthe intake stroke period and the spark-ignition is performed near theCTDC. In the first cycle, the air-fuel ratio of the mixture gas is setto the theoretical air-fuel ratio (λ≈1), and in order to adjust thefresh air amount to meet the fuel injection amount, the intake VVL 73controls the intake valve 21 to be operated with the large lift cam andthe VVT 72 sets the close timing of the intake valve 21 to a late timingafter the IBDC. By closing the intake valve 21 at the late timing, thefresh air amount is regulated (see the gas state inside the cylinderillustrated in the lowest row of FIG. 7). Moreover, in the example ofFIG. 7, in the first cycle, the amount of fresh air cannot be regulatedsufficiently by the control of the intake valve 21 alone and iscompensated by throttling the throttle valve 36. Note that the throttlevalve 36 is gradually opened further to prepare for switching to the CImode where the throttle valve is set to be fully opened. Moreover, inthe SI mode in the first cycle of FIG. 7, the external EGR gas is notintroduced as described above. Furthermore, the exhaust VVL 71 is turnedoff, in other words, the internal EGR gas is also not introduced. Thus,in the first cycle where the spark-ignition combustion is performed, theexhaust gas temperature becomes high (high-temperature burned gas), butboth the external and internal EGR gases are not introduced into thecylinder 18. Therefore, in the following second cycle, the exhaust gasis substantially not introduced into the cylinder 18.

The second cycle corresponds to a cycle when switching from the SI modeto the CI mode, and also corresponds to a transition mode. Here, theoperation of the ignition plug 25 is stopped to perform thecompression-ignition combustion. Moreover, in the second cycle, thethrottle valve is fully opened and the intake VVL 73 switches theoperating cam from the large lift cam to the small lift cam. Note thatthe VVT 72 is not actuated here, and the phase of the intake valve 21 isnot changed. Thus, the close timing of the intake valve 21 is instantlyswitched from the retarded timing in the first cycle to near the IBDC,and as a result, the amount of fresh air introduced into the cylinder 18increases. Note that the open and close timings of the intake valve 21here correspond to the timings in the exhaust open-twice controldescribed later.

Moreover, in the second cycle, the timing of the injection by theinjector 67 is set to be after the middle stage of the compressionstroke, and the fuel injection amount is increased compared to that inthe first cycle. Since the fresh air amount is increased, in the secondcycle, the A/F of the mixture gas is maintained substantially at thetheoretical air-fuel ratio. Moreover, as described above, in the secondcycle, the exhaust gas (the external EGR gas and the internal EGR gas)is not introduced.

Thus, in the second cycle, since a large amount of fresh air withcomparatively low temperature is introduced into the cylinder 18 byincreasing the amount of fresh air introduced into the cylinder 18, thein-cylinder temperature is decreased. Moreover, since the fuel isinjected into the cylinder 18 after the middle stage of the compressionstroke, by the latent heat of vaporization of fuel, the in-cylindertemperature is further decreased. Further, since the fuel injectionamount is increased, the amount by which the in-cylinder temperature isdecreased by the latent heat of vaporization may further be increased.Moreover, since the fuel injection timing is retarded, the reactabletime length of the mixture gas is shortened. As a result, in the secondcycle which is the cycle following the first cycle corresponding to theSI mode and where the in-cylinder temperature easily becomescomparatively high, the compression-ignition combustion can be startedat a suitable timing. Due to the decreased in-cylinder temperature andthe compression-ignition combustion with high thermal efficiency, thecombusting gas temperature is decreased greatly and the exhaust gastemperature is decreased. Note that in the gas state illustrated in thelowest row of FIG. 7, the relative temperature of the “burned gas” isindicated by the pitch width of the hatching pattern, in which thenarrow pitch width indicates that the burned gas temperature is high andthe wide pitch width indicates that the burned gas temperature is low.

The following third cycle corresponds to the cycle immediately afterswitching from the SI mode to the CI mode. In the third cycle, theexhaust open-twice control is performed by turning the exhaust VVL 71on. Thus, a part of the burned gas produced by the compression-ignitioncombustion in the second cycle is introduced into the cylinder 18;however, the burned gas temperature is suppressed to a low temperatureas described above. Moreover, since the in-cylinder temperature in thesecond cycle is suppressed to a low temperature, the in-cylindertemperature in the third cycle does not become significantly high.Moreover, in the third cycle, since the intake valve 21 still has asmall lift and the throttle valve 36 is fully opened, the same as in thesecond cycle, as illustrated in FIG. 7, the fresh air amount is reducedby the amount of the internal EGR gas introduced into the cylinder 18.

Since the third cycle corresponds to the normal CI mode, the timing ofthe injection by the injector 67 is set to be between the intake strokeand the early stage of the compression stroke, and the fuel injectionamount is reduced compared to that in the second cycle. Thus, the A/F ofthe mixture gas in the third cycle is set to the theoretical air-fuelratio or leaner than the theoretical air-fuel ratio according to theoperating state of the engine 1.

Also in the third cycle, similarly to the second cycle, the ignitionplug 25 is not actuated. As described above, in the third cycle, sincethe in-cylinder temperature does not increase excessively, by injectingthe fuel during the intake stroke, the comparatively homogeneous mixturegas formed within the cylinder 18 is surely compressed to ignite nearthe CTDC and stably combusts without causing pre-ignition. Thus, theincrease in combustion noises immediately after switching from the SImode to the CI mode is avoided. The switch from the SI mode to the CImode is completed as described above, and therefore, after the thirdcycle, a combustion control according to the operating state of theengine 1 is performed.

FIG. 8 is a time chart for a transition control performed when switchingfrom the part of the SI mode where the external EGR gas is introduced,to the CI mode. This time chart corresponds to, for example, a situationof switching from the high engine load range of the SI mode to the lowengine load range of the CI mode. FIG. 8 illustrates one example of thetransition from the range (3) (or the range (4)) in the SI mode to therange (1) (or the range (2)) in the CI mode in the operation map in thewarmed-up state illustrated in FIG. 4.

In other words, in the first cycle corresponding to the leftmost part inFIG. 8, the engine is operated in the SI mode, and here, the fuel isinjected in a period from the late stage of the compression stroke tothe early stage of the expansion stroke (i.e., the high pressureretarded injection) and the spark-ignition is performed near the CTDC.This corresponds to the ignition at the MBT. The air-fuel ratio of themixture gas is set to the theoretical air-fuel ratio (λ≈1), and in orderto adjust the fresh air amount to meet the fuel injection amount, theintake VVL 73 controls the intake valve 21 operated with the large liftcam and the VVT 72 sets the close timing of the intake valve 21 to alate timing after the IBDC. By closing the intake valve 21 at the latetiming, the fresh air amount is regulated. Moreover, the first cycle inFIG. 8 is the same as the first cycle in FIG. 7 in the sense that thethrottle valve 36 is throttled but it is gradually opened further toprepare for switching to the CI mode. Whereas, in the first cycle inFIG. 8, as described above, the EGR valve 511 and/or the EGR coolerbypass valve 531 are opened to introduce the external EGR gas into thecylinder 18. Note that the exhaust VVL 71 is turned off and the internalEGR gas is not introduced. In the first cycle where the spark-ignitioncombustion is performed, the exhaust gas temperature may become high.

The following second cycle corresponds to a cycle when switching fromthe SI mode to the CI mode, and also corresponds to the transition mode.In the transition mode here, the spark-ignition combustion is performed.In other words, in the second cycle, the throttle valve is fully openedand the intake VVL 73 maintains the large lift cam. Moreover, the EGRvalve 511 and the EGR cooler bypass valve 531 are opened, and theexternal EGR gas is introduced into the cylinder 18. Therefore, the gasstate inside the cylinder is not substantially different from the firstcycle.

The amount of the fuel injected by the injector 67 is set approximatelythe same as that in the first cycle. Thus, in the second cycle, the A/Fof the mixture gas is set to the theoretical air-fuel ratio. Note that,as described later, the ignition timing is advanced earlier than in thefirst cycle, and therefore, the fuel injection timing is advancedearlier than in the first cycle where the high-pressure retardedinjection is set.

The ignition timing in the second cycle is advanced earlier than in thefirst cycle, in other words, advanced by more than the MBT. Thus, whilethe generated torque is substantially the same as that in the firstcycle, the combusting gas temperature is decreased, and as a result, thedischarged exhaust gas temperature is decreased. Here, for the sake ofconvenience of explanation, the burned gas is referred to as themid-temperature burned gas to indicate that the burned gas has atemperature around the middle between the high-temperature burned gasand low-temperature burned gas. In the following third cycle, if theexhaust VVL 73 is turned on and a large amount of mid-temperature burnedgas is introduced into the cylinder 18, the in-cylinder temperature maybecome excessively high in the compression-ignition combustion.

The following third cycle corresponds to the second cycle in the timechart of FIG. 7, and also corresponds to the transition mode. In otherwords, in the time chart of FIG. 8, the transition mode is performed,not only over one cycle, but over a plurality of cycles.

In the third cycle, the operation of the ignition plug 25 is stopped toperform the compression-ignition combustion. Moreover, in the thirdcycle, the throttle valve is kept fully open and the intake VVL 73switches the operating cam from the large lift cam to the small liftcam. Thus, the close timing of the intake valve 21 is instantly switchedto the early timing near the IBDC, and as a result, the amount of freshair introduced into the cylinder 18 is increased.

Also in the third cycle, the EGR valve 511 and the EGR cooler bypassvalve 531 are fully closed to stop the introduction of the external EGRgas into the cylinder 18. However, the external EGR has low controlresponsiveness. Therefore, even after the EGR valve 511 and the EGRcooler bypass valve 531 are fully closed, the mid-temperature exhaustgas remaining in the EGR passage 50 is introduced into the cylinder 18in the third cycle (see the gas state illustrated in the lowest row ofFIG. 8).

Further, in the third cycle, similarly to the description above, thetiming of the injection by the injector 67 is set to be after the middlestage of the compression stroke, and the fuel injection amount isincreased compared to that in the first and second cycles.

Thus, in the third cycle, as described above, by increasing the amountof fresh air with comparatively low temperature introduced into thecylinder 18, the in-cylinder temperature is decreased, and then byinjecting a large amount of fuel into the cylinder 18 after the middlestage of the compression stroke, the in-cylinder temperature is furtherdecreased by the latent heat of vaporization of the fuel. As a result,in the third cycle, the compression-ignition combustion can be startedat a suitable timing and, due to the decreased in-cylinder temperatureand the compression-ignition combustion with high thermal efficiency,the exhaust gas temperature can be decreased significantly (i.e.,low-temperature burned gas can be produced).

The fourth cycle corresponds to the cycle immediately after switchingfrom the SI mode to the CI mode, and also corresponds to the third cyclein the time chart of FIG. 7. In the fourth cycle, the exhaust VVL 71 isturned on to perform the exhaust open-twice control. Thus, part of theburned gas with low temperature produced by the compression-ignitioncombustion in the third cycle is introduced into the cylinder 18. As aresult, the in-cylinder temperature in the fourth cycle does not becomesignificantly high. Moreover, in the fourth cycle, since the intakevalve 21 still has a small lift and the throttle valve 36 is fullyopened, the fresh air amount is reduced by the amount of the internalEGR gas introduced into the cylinder 18.

Since the fourth cycle corresponds to the normal CI mode, the timing ofthe injection by the injector 67 is set to be between the intake strokeand the early stage of the compression stroke, and the fuel injectionamount is reduced compared to that in the third cycle corresponding tothe transition mode. Thus, the A/F of the mixture gas in the fourthcycle is set to the theoretical air-fuel ratio or leaner than thetheoretical air-fuel ratio according to the operating state of theengine 1.

In the fourth cycle, the compression-ignition combustion is performed bystopping the operation of the ignition plug 25. As described above,since the in-cylinder temperature is not excessively high, by injectingthe fuel during the intake stroke, the comparatively homogeneous mixturegas formed within the cylinder 18 is surely compressed to ignite nearthe CTDC and stably combusts without causing pre-ignition.

As illustrated in FIG. 8, for example, in the high engine load range ofthe SI mode, the in-cylinder temperature may increase due to theincrease in fuel amount, and the exhaust gas temperature may alsoincrease. Thus, pre-ignition may easily occur particularly whenswitching from the high engine load range of the SI mode to the lowengine load range of the CI mode; however, by performing the transitioncontrol over two cycles corresponding to the transition mode where thespark-ignition combustion is performed and the transition mode where thecompression-ignition combustion is performed as illustrated in FIG. 8,pre-ignition can surely be avoided.

Next, regarding the transition control described above, a control flowperformed by the PCM 10 is described with reference to FIG. 9. This flowstarts in the SI mode with the air excess ratio λ=1. At S91 after thestart of the flow, the PCM 10 reads the various parameters (e.g., theengine coolant temperature, an outdoor air temperature, the engine load,the engine speed, the fuel injection timing, the fuel pressure, theignition timing, the open and close timings of the intake valve, theopen and close timings of the exhaust valve) to grasp the operatingstate of the engine 1. Then, at S92, the PCM 10 determines whether theSI mode is to be switched to the CI mode. If the mode is not to beswitched (i.e., S92: NO), the processing of S91 and S92 is repeated,whereas if the mode is to be switched to the CI mode (i.e., S92: YES),the flow proceeds to S93. Specifically, as described above, the resultof the determination at S92 is YES when the engine 1 shifts from thecold-start to the warmed-up state, when the operating state of theengine 1 shifts from the idle state to the low engine load operatingstate other than the idle state, when the engine 1 shifts from the statewhere the SI mode is performed temporarily after resuming from the fuelcut, to the normal CI mode, etc. Moreover, the result of thedetermination is also YES when the engine load is decreased and theoperating range is shifted from the high engine load range of the SImode to the low engine load range of the CI mode. From the start of theflow to S92 corresponds to the first cycle in the time charts of FIGS. 7and 8.

At S93, it is determined whether the external EGR gas is introduced, andif the external EGR gas is introduced (i.e., S93: NO), the flow proceedsto S94, whereas if the external EGR gas is not introduced (i.e., S93:YES), the flow proceeds to S96. At S94, the ignition timing is advancedand the spark-ignition combustion is performed. S94 corresponds to thesecond cycle in the time chart of FIG. 8. At the following S95, both theEGR valve 511 and the EGR cooler bypass valve 531 are fully closed, andthen the flow proceeds to S96.

At S96, the intake VVL 73 switches the operating cam from the large liftcam to the small lift cam. Then at S97, the fuel injection timing is setto be after the middle stage of the compression stroke and the fuelinjection amount is increased, and then at S98, the compression-ignitioncombustion is performed. The processing from S96 to S98 corresponds tothe second cycle in the time chart of FIG. 7, and also corresponds tothe third cycle in the time chart of FIG. 8.

Further at S99, the exhaust VVL 71 is turned on to start the exhaustopen-twice control. Then, at S910, the fuel injection timing is changedto be during the intake stroke and the compression-ignition combustionis performed. Thus, the switch to the normal CI mode is complete. Theprocessing at S99 and 5910 corresponds to each of the third cycle inFIG. 7 and the fourth cylinder in FIG. 8.

Here, with the configuration described above, when switching from the SImode to the CI mode, the transition mode over a single cycle (FIG. 7) ortwo cycles (FIG. 8) intervenes; however, it may be such that it isdetermined whether the exhaust gas temperature is decreased to apredetermined level, and the transition mode is continued until theexhaust gas temperature reaches the predetermined level. The exhaust gastemperature may be estimated based on the various parameters read by thePCM 10, for example.

Moreover, it may be configured such that by setting in advance thenumber of cycles where the transition mode is performed and storing itin the PCM 10 based on, for example, the operating state of the engine 1slightly before or after the switching from the SI mode to the CI mode,the transition mode is continued over the set number of cycles whenswitching from the SI mode to the CI mode.

Further, in the time chart of FIG. 8, the transition mode corresponds tothe combination of the cycle where the spark-ignition combustion isperformed and the cycle where the compression-ignition combustion isperformed; however, the transition mode may only correspond to the cyclewhere the spark-ignition combustion is performed.

Note that with the configuration described above, the valve operatingmechanism of the intake valve 21 includes the VVL 73 which switchesbetween the large lift cam and the small lift cam. The valve operatingmechanism of the intake valve 21 may include, instead of the VVL, a CVVL(Continuous Variable Valve Lift) for continuously varying the lift. TheCVVL can suitably adopt various known structures (detailed structurethereof is not illustrated). With the VVT and the CVVL, the open andclose timings and the lift (and the open period) of the intake valve 21can continuously be varied.

FIGS. 10 and 11 are time charts illustrating the switch controls fromthe SI mode to the CI mode with the configuration in which the valveoperating mechanism of the intake valve 21 includes the CVVL. FIG. 10relates to the switch from the part of the SI mode where the externalEGR gas is not introduced, to the CI mode, which corresponds to FIG. 7.FIG. 11 relates to the switch from the part of the SI mode where theexternal EGR gas is introduced to the CI mode, which corresponds to FIG.8.

First, in the first cycle corresponding to the leftmost part in FIG. 10,the engine is operated in the SI mode, the air-fuel ratio of the mixturegas is set to the theoretical air-fuel ratio (λ≈1). In order to adjustthe fresh air amount to meet the fuel injection amount, the CVVL of theintake valve 21 operates the intake valve 21 with comparatively smalllift and the VVT 72 sets the close timing of the intake valve 21 to acomparatively early timing before the IBDC. By closing the intake valve21 at the early timing, the fresh air amount is regulated. Moreover, inthe first cycle, the amount of fresh air cannot be regulatedsufficiently by the control of the intake valve 21 alone and iscompensated by throttling the throttle valve 36. Note that the throttlevalve 36 is gradually opened further to prepare for switching to the CImode where the throttle valve is set to be fully opened. Moreover, inthe first cycle, the external EGR gas is not introduced as describedabove. Furthermore, the exhaust VVL 71 is turned off.

The second cycle corresponds to the transition mode, and here, the VVT72 is actuated to retard the close timing of the intake valve to nearthe IBDC. By this, the fresh air amount introduced into the cylinder 18is increased compared to that in the first cycle. Moreover, in thesecond cycle, the fuel injection timing is set to be after the middlestage of the compression stroke, and the fuel injection amount isincreased. Thus, in the second cycle, as described above, the decreaseof the in-cylinder temperature before the compression stroke byintroducing a large amount of fresh air into the cylinder 18, and thedecrease of the in-cylinder temperature at the end of the compressionstroke by injecting a large amount of fuel into the cylinder on thecompression stroke and using the latent heat of vaporization of the fuelare achieved, and the in-cylinder temperature after the compressionstroke is decreased significantly. Thus, in the second cycle, theoperation of the ignition plug 25 is stopped to perform thecompression-stroke combustion, and the temperature of the exhaust gasdischarged is decreased.

The third cycle corresponds to the cycle after switching to the CI mode.Here, the exhaust VVL 71 is turned on to start the exhaust open-twicecontrol. Thus, by introducing the low-temperature exhaust gas into thecylinder 18, the compression-ignition combustion is performed stably.

Next, the first cycle corresponding to the leftmost part in FIG. 11 isin the high engine load range of the SI mode where the external EGR gasis introduced, and the air-fuel ratio of the mixture gas is set to thetheoretical air-fuel ratio (λ≈1). In order to adjust the fresh airamount to meet the fuel injection amount, the CVVL of the intake valve21 operates the intake valve 21 with comparatively small lift and theVVT 72 sets the close timing of the intake valve 21 to a comparativelyearly timing before the IBDC. Thus, the fresh air amount is regulated.Moreover, in the first cycle, the external EGR gas is introduced intothe cylinder 18. Furthermore, the exhaust VVL 71 is turned off.

The second cycle corresponds to the transition mode with thespark-ignition combustion. In the second cycle, the close timing of theintake valve 21 is not changed but the ignition timing is advanced bymore than the MBT. Note that the fuel injection timing is also advanceddue to the advance of the ignition timing. Thus, the combusting gastemperature is decreased to produce the mid-temperature exhaust gaswithout reducing the torque.

The third cycle corresponds to the transition mode with thecompression-ignition combustion. Here, both the EGR valve 511 and theEGR cooler bypass valve 531 are closed and the operation of the ignitionplug 25 is stopped. Note that, in this cycle, a part of the burned gasremaining in the EGR passage 50 is introduced into the cylinder 18.Moreover, the close timing of the intake valve 21 is retarded to nearthe IBDC to increase the fresh air amount. Further, the timing of thefuel injection by the injector 67 is set to be after the middle stage ofthe compression stroke and the fuel injection amount is increased.

Thus, in the third cycle, the in-cylinder temperature is decreased bythe increase in the amount of fresh air introduced into the cylinder 18,the in-cylinder temperature is decreased by injecting the fuel into thecylinder 18 after the middle stage of the compression stroke, and thein-cylinder temperature is further decreased by the increase in the fuelinjection amount. Therefore, the compression-ignition combustion can bestarted at a suitable timing. As a result, the exhaust gas temperatureis decreased.

In the following fourth cycle, the exhaust VVL 71 is turned on to startthe exhaust open-twice control, and by introducing the low-temperatureexhaust gas, the compression-ignition combustion can be performed stablywithout causing pre-ignition.

As described above, also with the valve operating mechanism of theintake valve 21 including the CVVL, a similar transition control can beperformed. Note that, the VVL 73 can instantly switch the intake mannerin the transition control, and is excellent in view of improvingresponsiveness of the transition control and switching of the modesmoothly.

Note that the application of the art disclosed herein is not limited tothe engine configuration described above. For example, the fuelinjection in the intake stroke period may be performed into the intakeport 16 by a port injector separately provided in the intake port 16,instead of the injector 67 provided in the cylinder 18.

Moreover, the engine 1 is not limited to the in-line four cylinderengine described above, and may be applied to an in-line three cylinderengine, an in-line two cylinder engine, an in-line six cylinder engine,etc. Further, the engine 1 is applicable to various kinds of engines,such as a V6 engine, a V8 engine, and a flat-four engine.

Moreover, the operating ranges illustrated in FIG. 4 are merely anexample, and other various operating ranges may be provided.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS

-   1 Engine (Engine Body)-   10 PCM (Controller)-   18 Cylinder-   25 Ignition Plug-   67 Injector (Fuel Injection Valve)

What is claimed is:
 1. A control device of a spark-ignition engine,comprising: a main body of the engine having a cylinder; a fuelinjection valve for injecting fuel supplied into the cylinder; anignition plug for igniting mixture gas within the cylinder; and acontroller for operating the engine by controlling at least the fuelinjection valve and the ignition plug, wherein according to an operatingstate of the engine, the controller switches an ignition mode between acompression-ignition mode in which compression-ignition combustion isperformed by causing the mixture gas to self-ignite and combust, and aspark-ignition mode in which spark-ignition combustion is performed byigniting the mixture gas with the ignition plug to combust the mixturegas, and wherein the controller switches the ignition mode from thespark-ignition mode to the compression-ignition mode by performing atransition mode in which a temperature inside the cylinder is forciblydecreased and combustion is performed.
 2. The device of claim 1, whereinthe controller decreases the temperature inside the cylinder andperforms the compression-ignition combustion by injecting the fuel withthe fuel injection valve from an intake stroke to an early stage ofcompression stroke in the compression-ignition mode, and at least aftera middle stage of the compression stroke in the transition mode.
 3. Thedevice of claim 2, wherein the controller increases an amount of freshair introduced into the cylinder in the transition mode.
 4. The deviceof claim 3, wherein in the transition mode, the controller increases anamount of the fuel injected into the cylinder, corresponding to theincrease in the fresh air amount.
 5. The device of claim 1, wherein inthe transition mode, the controller performs the spark-ignitioncombustion and decreases a temperature of combusting gas by advancing anignition timing of the ignition plug by at least a minimum advance forbest torque.
 6. The device of claim 5, further comprising an externalexhaust gas recirculation (EGR) adjuster for circulating exhaust gasdischarged from the cylinder back to the intake side, wherein whenswitching from the spark-ignition mode where the external EGR adjusteris operated to circulate the exhaust gas into the cylinder to thecompression-ignition mode, in the transition mode, the controllerperforms the spark-ignition combustion and advances the ignition timingof the ignition plug by at least the minimum advance for best torque. 7.The device of claim 1, further comprising an external EGR adjuster forcirculating exhaust gas discharged from the cylinder back to the intakeside, wherein the controller performs the compression-ignitioncombustion in the transition mode when switching from the spark-ignitionmode where the external EGR adjuster is not operated, to thecompression-ignition mode.
 8. The device of claim 7, wherein when theengine load is at a predetermined high load, the controller controls theignition mode to the spark-ignition mode where the external EGR adjusteris operated, and when the engine load is at a predetermined low load,the controller controls the ignition mode to the spark-ignition modewhere the external EGR adjuster is not operated.